Engineers typically design high-pressure oil field plunger pumps in two sections: the (proximal) power section and the (distal) fluid section. The power section usually comprises a crankshaft, reduction gears, bearings, connecting rods, crossheads, crosshead extension rods, etc. Power and fluid sections are commonly referred to in the industry, and hereafter in the application, as the power end and fluid end, respectively. Fluid ends usually comprise a plunger pump fluid end housing with multiple internal cavities or fluid chambers, each chamber having a suction valve in a suction bore, a discharge valve in a discharge bore, and a plunger in a plunger bore, plus high-pressure seals, retainers, etc. FIG. 1 is a cross-sectional schematic view of a typical fluid end housing showing its connection to a power end by stay rods. A plurality of fluid chambers similar to that illustrated in FIG. 1 may be combined, as suggested in the Triplex fluid end housing comprising three (3) fluid chambers is schematically illustrated in FIG. 2. A pump with five (5) fluid chambers or 5 plungers is referred to as a quintuplex pump.
Valve terminology varies according to the industry, e.g., pipeline or oil field service, in which the valve is used. In some applications, the term “valve” means just the moving element or valve body. In the present application, however, the term “valve” includes other components in addition to the valve body, e.g., various valve guides to control the motion of the valve body, the valve seat, and/or one or more valve springs that tend to hold the valve closed, with the valve body reversibly sealed against the valve seat.
Valve and seat sizing design is a compromise between competing objectives in fluid end design. Traditionally engineers have wanted to use suction valve and seat designs of as a large a size as possible, as the larger the flow area in the valve and seat, the lesser the flow restriction. Flow restrictions reduce fluid energy which hinders the complete filling of the fluid chamber and the volumetric efficiency of the pump. Incomplete filling of the fluid chamber can cause a rough running pump. Additionally, larger valve and seat sizes reduce fluid velocity through the valve and seats. High fluid velocity contributes to erosion damage of the valve seal and leads to premature seal failure of the valve. For additional detail on valve erosion damage read the teaching of U.S. Pat. No. 9,416,887. The disadvantage of larger valve and seat sizes is the greater the size and weight of the fluid end housing necessary to contain the larger size valve and seat. Larger valve and seat sizes also result in higher valve loads and higher stresses on the fluid end housing which can result in premature structural failure of the housing. In rare instances in the prior art, suction valves and seats were slightly larger than discharge valves and seats. The theoretical reason for this sizing was based on the belief that greater flow area was necessary in the suction valves and seats to reduce flow restrictions than comprised fluid energy in filling the fluid chamber on the suction stroke. Further, many designers observed that the fluid in the discharge stroke inherited great fluid energy from the applied power of the moving plunger and thus smaller valve and seat sizing could be applied to the discharge valves and seats. This reasoning ignores the requirement to reduce fluid velocity in both sets of valves and seats to prevent erosion damage and premature failure to valve seals.
Similarly in the prior art, the suction port and discharge were almost always maximized to reduce flow restrictions. The suction port and discharge port are the volumetric bores directly upstream and feeding the suction valve/seat and discharge valve/seat, respectively. The respective bore of these respective ports would typically be maximized by boring the port to the small diameter of the taper in the fluid end housing utilized in capturing and securing the suction or discharge seat. This design practice was justified in the suction port because of the need to preserve fluid energy by reducing flow restrictions. By default, the same practice was utilized for the discharge port. As will be discussed later in this application, a large discharge port is not warranted.
Each individual bore in a plunger pump fluid end housing is subject to fatigue due to alternating high and low pressures which occur with each stroke of the plunger cycle. Conventional fluid end housings, also referred to as Cross-Bore blocks, typically fail due to fatigue cracks in one of the areas defined by the intersecting suction, plunger, access and discharge bores as schematically illustrated in FIGS. 3A and 3B.
To reduce the likelihood of fatigue cracking in the high-pressure plunger pump fluid end housings described above, a Y-block housing design has been proposed. The Y-block housing design, which is schematically illustrated in FIG. 4A, reduces stress concentrations in a plunger pump housing, such as that shown in FIG. 3A, by increasing the angles of bore intersections above 90°. In the illustrated example of FIG. 4A, the bore intersection angles are approximately 120°. A more complete cross-sectional view of a Y-block plunger pump fluid end housing and the assembly components is schematically illustrated in FIG. 4B.
Both cross-bore blocks and Y-blocks have several major disadvantages when used to pump heavy slurry fluids as typically utilized in oilfield fracturing service. A first disadvantage is related to the feeding of the fluid chamber on the suction stroke of the pump. Upon passing through the suction valve, the fluid must make a 90 degree turn in a cross-bore housing, or a 60 degree turn in a Y-block housing, into the plunger bore as illustrated in FIG. 5. This change in the direction of the heavy fluid robs the fluid of kinetic energy, hereafter referred to as fluid energy.
Fluid energy is normally added to the fluid by small supercharging pumps upstream from the plunger pump. Fluid energy is necessary to overcome fluid inertia and ensure complete filling of the fluid chamber on the suction stroke. If the fluid could enter the fluid chamber in a linear or straight path, less fluid energy would be lost.
The second disadvantage of Cross-Bore blocks and Y-blocks relates to the large intersecting curved areas where the various bores intersect. Because the suction bore above the suction valve is almost as large as the plunger bore, the intersection area of the suction bore with the plunger bore is particularly large, as illustrated in FIGS. 3A and 3B. While the intersection of the suction bore and the plunger bore is especially large, the intersection of the discharge bore and the plunger bore is almost as large.
As shown in FIGS. 6A and 6B, the intersecting cylindrical sections result in intersection curves that focus or concentrate the stresses generated by the internal pump pressures into a very small area. This small area is located at the bore intersection near the plane formed by the centerline axis of the plunger and suction or discharge bore cylinders at the finite point of the intersection of the two cylinders. Because the intersection curve changes slope through three-dimensional space, this intersection cannot be easily chamfered or filleted by conventional machining techniques that would mitigate these stresses. Indeed, complex computer finite element stress analysis calculations indicate that chamfering or filleting the corner intersection has minimal effect on reducing the stresses at this corner intersection.
The amount of stress at the intersecting bores of conventional fluid end housings is defined by the magnitude of the “Bore Intersection Pitch” as illustrated in FIGS. 3A, 3B, and 4A. Any geometry that reduces the “Bore Intersection Pitch” will reduce the stress concentrations in the fluid end and increase the life of the fluid end by mitigating cyclic fatigue failure. Y-Block fluid end housing designs, such as those illustrated in FIG. 4A, do reduce this pitch, but the reduction is insufficient to prevent cyclic fatigue failure of the fluid end housing when subjected to high pressure and long pumping cycles.
Previously filed U.S. Non-Provisional patent application Ser. No. 15/330,212, filed on Aug. 23, 2016, and U.S. Non-Provisional patent application Ser. No. 15/330,213, filed on Aug. 23, 2016, featured an “in-line” design and addressed many of the issues of failure due to high stress and “Bore Intersection Pitch.” These applications also addressed the loss of fluid energy at the intersection of the suction bore and plunger bore in typical cross bore designs illustrated in FIGS. 1, 2, 3A, 3B, 4A, 4B, and 5 of those patent applications.
One of the major shortcomings of the U.S. application Ser. Nos. 15/330,212 and 15/330,213 relates to maintenance complications encountered when changing the plunger or plunger packing. Fluid ends built to Ser. Nos. 15/330,212 and 15/330,213 require removal of the entire fluid end assembly to access the damaged or worn parts. This problem could be addressed with a two-piece plunger design; however, such plungers are difficult to access for maintenance and are prone to premature failure. A design similar to that disclosed in prior art application Ser. No. 15/330,212, with a modification to allow access for maintenance to plungers, packing, suction valve, and the suction seat would provide a major and much needed improvement.